It is briefly recalled that in a set of valves controlling respective loads that require different hydraulic powers, the load sensing system consists in detecting which one of the loads requires the maximum power and thus the maximum pressure in the working hydraulic fluid fed thereto, and in applying said maximum pressure to a control inlet of the pump so as to servo-control the pump to requirements. This function is implemented by providing each control valve with a selector that is responsive on one side to the pressure of the working fluid delivered to the load controlled by the valve and on its opposite side to the pressure of the working fluid delivered to another load controlled by a control valve and which is suitable for selecting the higher of said two pressures. By performing stepwise selection, it is the maximum pressure of the entire hydraulic system that finally controls the pump.
Installing the means (selectors and link channels) required for implementing a load sensing system within the control valves gives rise to a control valve structure that is quite complex. Various simplifications have been found for certain types of control valve, but none has yet been found for directional control valves that operate proportionally.
In addition, the load sensing lines are conventionally fed from a pressure take-off point formed at the load. When hydraulic fluid is first delivered, the load sensing line is fed with fluid before the load itself is. If the sensing line has a leak (and such a leak may be provided deliberately in certain modes of operating hydraulic circuits), the control pressure applied to the load begins by decreasing before it increases to the nominal value imposed by the control valve. As a result, the load (e.g. a hinged arm) begins by moving down before it moves up in compliance with the control applied thereto, and in any event a jolt occurs at the instant at which normal conditions are re-established. That constitutes a real drawback of the system which may turn out to be dangerous.
Furthermore, in a conventional hydraulic directional control valve, the hydraulic fluid flow rate delivered by the working orifice of the valve is subjected to fluctuations as a function of the magnitude of the flow rate as determined by the position of the slide and as a function of the pressure delivered by the pump. It is known that this drawback can be mitigated and the working fluid flow rate can be made constant regardless of circumstances (e.g. from U.S. Pat. No. 3,827,453) by providing pressure compensating means in the control valve that continuously compare the working pressure from the pump with a reference value that may be fixed or variable. If variable, it may be constituted by the maximum pressure as selected in the load sensing line, so as to throttle the working fluid accordingly, thereby establishing a constant pressure drop in said working fluid.
In known control valves (e.g. U.S. Pat. No. 4,693,272), the presence of such pressure compensating means further increases the complexity of the structure since although said pressure compensating means use the maximum pressure information present in the load sensing line, they are established independently of the means used for selecting the maximum pressure.
In addition, using such pressure compensation requires, in particular, a fraction of the necessary hydraulic links to pass through the slide. Drilling the corresponding ducts in the slide considerably increases the cost of manufacturing it. Furthermore, the presence of such ducts drilled through the slide occupies the internal volume thereof and it is no longer possible to provide other drillings that may be useful for other purposes, e.g. those required for implementing a load braking system. Such other systems then need to be designed in the form of circuits including external pipework, thereby further increasing complexity and expense of the assembly as a whole.
In other known directional control valves (U.S. Pat. No. 5,138,837, EP 0 438 606), attempts at simplifying and integrating the pressure compensating means and the load sensing means can indeed be found. However, the load sensing means continue to be implemented with a pressure take-off point situated in the line connected to the load: such known control valves therefore continue to suffer from the drawbacks mentioned above for that kind of organization.
It may also be added that in known directional control valves in which the pressure compensating function is provided by a spring-biased non-return valve, the pump-controlling pressure differs from the pressure of the pressurized fluid delivered by the pump not only by the pressure drop imposed by the pressure compensating means, but also by the head loss which is introduced by the non-return function provided by the non-return valve in the most heavily loaded control valve, corresponding to the rated value of the spring biasing the non-return valve. Thus, with such an organization, the presence of the return spring disturbs the ideal operation of the system, and this turns out to be a considerable drawback which makes itself felt most particularly in very low pressure ranges.